Refrigeration system with fluid transformer for controlling regrigerant flow

ABSTRACT

By connecting the primary and secondary flow circuits of a fluid transformer in different portions of a vapor compression refrigeration cycle, the refrigerant flowing in the primary may be used to control the refrigerant in the secondary and vice versa. In the disclosed embodiment, condensed liquid refrigerant expands in and drives the primary which in turn transfers energy to the secondary circuit as compression work to effect pumping of gaseous refrigerant through that circuit. A constant volumetric flow ratio is maintained between the flow rates in the two circuits, permitting the transformer to serve as a metering device to automatically meter the refrigerant flow in the system under wide ranges of evaporator heat loading and condenser temperature variation. During very low ambient condenser temperatures, giving rise to correspondingly low condenser head pressures, the primary and secondary flow circuits interchange roles with the gaseous refrigerant in the secondary now effecting pumping of the liquid refrigerant through the primary to the evaporator.

United States Patent Maudlin [l 1] 3,7 10,586 1451- Jan. 16, 1973 [s41 REFRIGERATION SYSTEM WITH FLUIDITRANSFORMER FOR CONTROLLING REGRIGERANT FLOW Primary Examiner-Meyer Perlin AttorneyDonald W. Banner,Wi11iam S. McCurry and John W. Butcher Secondary Inlet Secondary Outlet Fluid Primary Suction Line l6 Transformer o pl'll'llol'y Line 57 ABSTRACT By connecting the primary and secondary flow circuits of a fluid transformer in different portions of a vapor compression refrigeration cycle, the refrigerant flowing in the primary may be used to control the refrigerant in the secondary and vice versa. In the disclosed embodiment, condensed liquid refrigerant expands in and drives the primary which in turn transfers energy to the secondary circuit as compression work to effect pumping of gaseous refrigerant through that circuit. A constant volumetric flow ratio is maintained between the flow rates in the two circuits, permitting the transformer to serve as a metering device to automatically meter the refrigerant flow in the system under wide ranges of evaporator heat loading and condenser temperature variation. During very low ambient condenser temperatures, giving rise to correspondingly low condenser head pressures, theprimary and secondary flow circuits interchange roles with the gaseous refrigerant in the secondary now effecting pumping of the liquid refrigerant through the primary to the evaporator.

9 Claims, 5 Drawing Figures Discharge Line Condenser IE REFRIGERATION SYSTEM WITH FLUID TRANSFORMER FOR CONTROLLING REGRIGERANT FLOW BACKGROUND OF THE INVENTION This invention relates to a novel flow control arrangement for the refrigerant in a refrigeration system. It may be employed in a variety of different systems and in accordance with many different modifications. While the invention is susceptible to widely diverse uses in the field of refrigeration, it is particularly useful in all-weather air conditioning equipment required to operate in the presence of a broad range of outside ambient temperatures, and will be described in that en vironment.

The condenser of an air conditioning system is usually air cooled and is located outside. Since the condensing head pressure is directly proportional to the outside ambient temperature, the pressure of the condensed refrigerant in an all-weather system (one that must provide cooling all year round) will be substantially lower in the winter as compared to its summer level. For example, in a typical air conditioning system, employing R-22 for the refrigerant, the pressure of that refrigerant at the condenser outlet will be around 300 psig when the outdoor temperature is 95 F. At an outdoor temperature of F, on the other hand, the refrigerant leaving the condenser will have a pressure in the vicinity of 65 psig. A refrigeration system cannot function properly unless the condensing pressure exceeds a predetermined minimum level, customarily in the neighborhood of 170 psig which is that prevailing when the outside ambient is about 60 F. When the condensing pressure is below the required minimum, the expansion device cannot adequately feed the evaporator, thereby starving it and causing a lower evaporator temperature. The lower the ambient temperature the greater will be the amount of starving. Loss of capacity, poor coefficient of performance, and freeze up are the results. In addition, the low condensing head pressure under low ambient conditions reduces the saturation temperature in the liquid line as a consequence of which flash gas forms from heat inflow.

Previous efforts to solve this problem have generally involved the practice of artificially holding the condensing pressure up during low ambient operation in order to have sufficient liquid line pressure ahead or upstream of the expansion device. The artificially high condensing pressure, however, reduces the available coefficient of performance, thereby causing higher power costs than necessary. Attempts have been made in the past to reduce these power costs by employing outdoor air for cooling under low ambient conditions, but such systems are complex and expensive, requiring specially built machines having complicated control systems, multiple fans, dampers, etc.

Applicants unique, multi-purpose flow control scheme readily lends itself to incorporation in a refrigeration system to remedy the low ambient problem. The present invention permits a refrigeration system to operate under an extremely broad range of outdoor ambient temperature conditions, while at the same time providing the highest attainable coefficient of performance for the particular equipment in the system. Moreover, the refrigeration system is relatively inexpensive by comparison with the prior approaches and eliminates the need for using outdoor air for cooling during low ambient temperatures. In addition, applicants flow control arrangement provides stable liquid refrigerant metering to the evaporator under all conditions, maintaining the evaporator full of boiling liquid at all times. Furthermore, the invention provides fast liquid feed to the evaporator under cold start-up conditions, provides fast liquid feed shutoff following compressor shut down, and limits compressor loading at high ambients by limiting the suction pressure. Moreover, the invention does not require high refrigerant charges as is necessary in liquid back-up systems in order for the condensers to produce artificially high condenser pressures.

It is, therefore, an object of the invention to provide a new and improved refrigerant flow control arrangement for a refrigeration system.

Another object is to provide a novel refrigeration system that will function efficiently at extremely low condenser pressures.

A further object of the invention is to provide a novel flow control scheme for conveying refrigerant from a condenser to an evaporator during low ambient conditions and without artificially raising the condensing pressure.

It is still another object to provide a unique metering device for automatically metering the refrigerant flow in a refrigeration system subject to widely changing evaporator heat loads and condensing temperatures.

SUMMARY OF THE INVENTION A refrigeration system, constructed in accordance with a broad aspect of the invention, comprises a fluid transformer for utilizing (in its primary flow circuit) the refrigerant flowing in one portion of a vapor compression refrigeration cycle to control (in its secondary flow circuit) the refrigerant in another portion of the cycle.

In accordance with a narrower aspect, efficient operation under widely varying ambient conditions is achieved by connecting the primary flow circuit between the outlet of the condenser and the inlet of the evaporator in the compression cycle, and by connecting the secondary flow circuit between at least part of the evaporator and the inlet of the compressor. As long as the condenser head pressure is above a predetermined minimum level, liquid refrigerant in the primary circuit expands and in so doing transfers energy to the secondary to compress and pump gaseous refrigerant. Under conditions of low condenser pressure (below the required minimum level) the primary and secondary effectively interchange roles, the gaseous refrigerant expanding in the secondary and transfering energy to the primary to achieve pumping of the liquid refrigerant to a higher pressure and in the direction of the evaporator.

DESCRIPTION OF THE DRAWINGS The features of the invention which are believed to be novel are set forth with particularity in the appended claims. The invention, together with further objects and advantages thereof, may best be understood, however, by reference to the following description in conjunction with the accompanying drawings in which like reference numbers identify like elements, and in which:

FIG. I is a schematic representation of a refrigeration system constructed in accordance with one embodiment of the invention and including a fluid transformer for controlling refrigerant flow in the system;

FIGS. 2 and 3 each depicts the detailed construction of a preferred form of the fluid transformer and together those figures show the transformers operating cycle at four different steps one quarter of a cycle apart;

FIG. 4 is a pressure-enthalpy diagram illustrating the operation of the refrigeration system under high ambient conditions; and,

FIG. 5 is a pressure-enthalpy diagram indicating the operation of the system during low ambient conditions.

DESCRIPTION OF THE ILLUSTRATED EMBODIMENT Fluid transformer in the system of FIG. 1 may take any of a variety of different constructions. Its preferred form, shown in FIGS. 2 and 3, will be described in detail later. It is necessary, however, to be familiar with the basic characteristics and function of a fluid transformer in order that the operation of the il lustrated refrigeration system may be fully understood. Briefly, a transformer of the fluid type is a piston-actuated device having two isolated but interdependent flow circuits, the fluid (liquid or gas) passing through one of them (called the primary) normally controlling or driving the fluid in the other or secondary circuit. The primary will serve as the prime mover in the transformer so long as the pressure of the fluid at the inlet to the primary is greater than that of the fluid at the outlet. This means that the fluid expands as it flows through the primary. The pressure differential across the primary circuit in combination with the flow rate, namely the volume of fluid flowing through the primary per unit of time, produce energy which in turn is transmitted to the secondary flow circuit in the form of compression work to achieve pumping of the fluid through the secondary, the fluid compressing in the secondary and thus having a pressure at the secondary outlet that is higher than its pressure at the secondary inlet.

As is also usually characteristic of a fluid transformer, each of its primary and secondary flow circuits displaces a different volume of fluid per same unit of time and through at least one chamber. A constant volumetric flow ratio will exist between the two circuits as determined by the ratio of the two volumes displaced, namely the ratio between the total effective volume of the chamber or chambers of the primary circuit and the total effective volume of the chamber or chambers of the secondary circuit. In other words, the volume of fluid flowing through one of the flow circuits during a given time interval is directly proportional to the volume of fluid flowing through the other circuit during the same interval and vice versa. In this way, any variation in the flow rate in either one of the flow circuits results in a corresponding flow rate change in the other circuit in accordance with the volumetric flow ratio.

To explain further, a fluid transformer is somewhat analogous to an electrical transformer which has two isolated but interdependent primary and secondary windings. A voltage applied across the primary winding causes current conduction through that winding and this in turn produces a voltage across, and causes current to flow through, the secondary winding. The pressure differentials across the primary and secondary of a fluid transformer may be likened to the voltages across the primary and secondary of an electrical transformer, and the fluid flow rates are analogous to the currents. The ratio of volumes displaced in the fluid transformer is akin to the turns ratio of an electrical transformer, namely the number of primary winding turns compared to the number of secondary winding turns. Here, however, the analogy must relate the primary-to-secondary turns ratio to the volume displaced in the secondary flow circuit compared to the volume displaced in the primary circuit.

In an electrical transformer the ratio between the voltage applied to the primary and the voltage developed in the secondary is equal to and determined by the turns ratio. On the other hand, the ratio between the primary and secondary current flow is equal to the reciprocal of the turns ratio, or the number of secondary turns relative to the number of primary turns. The ratio between the primary and secondary pressure differentials in a fluid transformer is equal to the ratio between the volume displaced in the secondary relative to that displaced in the primary, and the ratio of the primary flow rate compared to the secondary flow rate equals the ratio of the primary volume displacement versus the secondary volume displacement.

The power in or energy supplied to an electrical transformer with a resistive load is determined by the product of the primary voltage and current and, ignoring losses, must be equal to the power out which is determined by the secondary voltage multiplied by the secondary current. Likewise, in a fluid transformer the power in or delivered to the primary equals the primary pressure differential times the flow rate of the primary fluid and, ignoring losses, that input energy must equal the output energy, namely the product of the secondary pressure differential and the flow rate of the secondary fluid. As is true in both transformer types, any variation of the power out results in a corresponding variation of the power in. For example, if the primary and secondary voltages remain constant in an electrical transformer but the secondary current changes (due to a load variation on the secondary) the primary current must correspondingly change in order for the power input and power output to match. Likewise, in a fluid transformer if the primary and secondary pressure differentials remain unchanged while the secondary flow rate changes in one sense the primary flow rate must vary in the same sense to equalize the input and output energies.

At least one movable piston member is generally employed in each of the chambers of the primary and secondary flow circuits in order to displace or move the fluid through those circuits. The transformer effectively has an operating speed which is directly proportional to the speed at which the piston or pistons of the primary move and this in turn is directly proportional to the flow rate of the driving fluid in the primary. The flow rate of the fluid flow in the secondary is directly proportional to the transformer's operating speed.

While in accordance with the invention there are a variety of different manners in which a fluid transformer may be incorporated in a refrigeration system to perform various diverse functions and to accomplish different desired results, in the particular environment chosen for illustration the primary circuit of fluid transformer is interposed or inserted in the vapor compression refrigeration cycle of an air conditioning system between the outlet of condenser 12 and the inlet of evaporator 14. In other words, the primary is connected in series with the liquid line from the condenser. The secondary flow circuit of the transformer is connected in series with the portion of the compression cycle between the outlet of evaporator 14 and the inlet of condenser 12. Specifically, it is connected in series with the suction line coupling the evaporator to the suction inlet of compressor 16 It will be observed that there is no separate expansion device which is normally included in the vapor compression refrigeration cycle between the outlet of the condenser and the inlet of the evaporator. Since the primary fluid must expand in the primary circuit in order to drive or pump fluid through the secondary, the primary circuit itself serves as the required expansion device for the compression cycle.

Assuming that refrigerant of the R-22 type is used in the system, transformer 10 will be constructed so that its pistons displace a volume in the secondary that is 45 times the volume displaced in the primary. Such a volumetric ratio is selected since an air conditioning system is customarily rated under conditions of 95 F dry bulb outdoor air and 67 F wet bulb indoor air, and with slightly superheated gaseous refrigerant emerging from the evaporator; and in the presense of those conditions the volume ratio between the gaseous and liquid states of refrigerant R-22 is 45 to l. Stated differently, one cubic foot of liquid converts to 45 cubic feet of gas. Hence, with a fluid transformer having a primary displacement one forty-fifth of its secondary displacement, liquid will be metered or fed to the evaporator at the rate of 1 cubic foot for every 45 cubic feet of suction gas flowing through the secondary to compressor 16, and thus the system is in balance. Under the rated conditions the fluid transformer will have a predeterminedoperating speed determined by the flow rate of the liquid refrigerant in the primary.

More particularly, refrigerant gas is received at the compressors suction inlet at a relatively low pressure and temperature and is compressed by the compressor in order to discharge that gas to the inlet of condenser 12 at a relatively high pressure (the high side pressure of the system) and temperature. Heat is removed from the hot discharge gas in the condenser while retaining essentially the same high side pressure. Sufficient heat is deleted to condense the discharge gas so that all of the refrigerant leaves condenser 12 in its liquid state. The liquid refrigerant in passing through the primary circuit of the fluid transformer expands and thus reduces its pressure (to the low side pressure of the system) and temperature, emerging at the primary outlet of the transformer as a mixture of liquid and gas but primarily a liquid. As the mixture then flows through evaporator 14, which is in heat exchange relation or contact with the medium to be cooled, a constant pressure is maintained while heat is transferred from the medium to the refrigerant and the entirety of the refrigerant assumes its gaseous state. Actually, sufficient liquid would be metered to the evaporator to fill it with boiling liquid and thereby maximize the amount of heat that can be absorbed by the refrigerant in the evaporator, while adding a relatively small desired amount of superheat to the refrigerant before it emerges at the evaporators outlet. This superheated gas is returned to the suction inlet of compressor 16 but first passes through the secondary circuit of transformer 10 wherein its pressure increases due to compression work transferred to the secondary as a result of the expansion of the liquid in the primary.

Specifically, the expansion of the refrigerant in the primary (which constitutes an expansion engine) from the high side pressure to the low side pressure provides a pressure differential which in conjunction with the flow rate of the liquid through the primary creates PV or pressurevolume energy which is then used in the secondary circuit to pump and compress the suction gas so that it enters compressor 16 at a pressure higher than it otherwise would. in effect, the transformer provides an additional compression stage. As a con sequence, compressor 16 does not have to compress the suction gas all the way from the low side to the high side pressure.

Moreover, by transfering energy from the primary to the secondary an improvement is gained in the performance over conventional systems inasmuch as the evaporator will be fed with a refrigerant having a lower heat content or enthalpy. This advantage is readily appreciated by referring to the pressure-enthalpy diagram of FIG. 4 which illustrates the operation of the system during high ambient conditions. The dashed construction lines in FlG. 4 indicate the path that is followed by a conventional refrigeration system of the prior art. In transfering PV energy to the secondary, the liquid refrigerant sub-cools and loses heat content in the amount AH As shown in FIG. 4, the refrigerant thus arrives at the evaporator with a much lower heat content. This means that more heat can be absorbed by the refrigerant from the medium to be cooled. The initial portion of the compression phase shown in FIG. 4 (starting from the low side pressure) represents the compression introduced in the secondary. That energy is denoted AH, and it is, of course, equal to AH,.

Assume now that the outdoor temperature suddenly increases above the rated conditions of F. The condenser head pressure will rise and this results in a higher discharge pressure at the outlet of the compressor. The pumping capacity of the compressor therefore decreases and this in turn reduces the quantity of gas (in terms of volume per unit of time) that can be removed from the evaporator and delivered through the secondary flow circuit to the compressors inlet. With a slower moving refrigerant in the evaporator more heat will be absorbed per pound circulated, as a consequence of which the refrigerant gas leaves the evaporator at a higher temperature and pressure. The higher pressure creates a lower specific volume (volume per pound) of the suction gas, and the lower the specific volume the lower will be the quantity of gas that will flow through the secondary. At an outdoor temperature of F, for example, 1 cubic foot of liquid refrigerant R-22 converts to only 41 cubic feet of gas, and this does not satisfy the fluid transformers requirement of 45 to 1 thereby causing the system to become unbalanced. In other words, if it is assumed that the outside temperature suddenly climbs from 95 F to 115 F every cubic foot of liquid passing through the primary creates only 41 cubic feet of gas through the secondary. In effect, there will be a 4 cubic foot void or unused space in the chamber or chambers of the secondary.

In response to the rising outdoor temperature, the transformer automatically reduces its operating speed and stabilizes at a new speed which will re-establish the required 45 to 1 ratio of gas volume versus liquid volume. To elucidate, with less gas to be pumped through the secondary less PV energy need be developed in the primary and transmitted to the secondary to perform the pumping work. With a lower energy requirement to be satisfied by the primary, the liquid flow (volume per unit of time) reduces and the transformer slows down, since its operating speed is directly proportional to the flow rate of the primary fluid. Such a response is consistent with that occurring in an electrical transformer when there is a decrease in the load requirement to be met by the secondary winding. At that time, the secondary current decreases and lowers the output power of the transformer. in order for the input power to match the lower power out, the primary current must decrease.

With transformer 10 functioning at a reduced speed and with a smaller quantity of liquid supplied to the evaporator, less heat can be absorbed from the heat load and this causes a drop in both the temperature and pressure of the gas exiting at the evaporators outlet. With a lower pressure, the suction gas volume per pound, or specific volume, increases. This action continues until the ratio of suction gas volume to liquid line volume once again becomes 45 to l. The net result is that, in response to higher condenser head pressures, the transformer runs slower than at its rated conditions and less pounds per minute of refrigerant are metered through the primary and into the evaporator, starving it in order to restrict the maximum suction pressure to that desirable for the particular compressor used in the system.

It is to be noted that the fluid transformer limits the pressure of the gas delivered to the compressor in much the same way as does a thermostatic expansion valve with the maximum operating pressure feature. However, the transformer will limit the suction pressure to that occurring at the equipment rating point while the thermostatic expansion valve cannot be applied this close to its maximum operating pressure setting.

When the outside temperature drops below the rated conditions, the condenser pressure reduces the ratio of suction gas volume flow to liquid line flow increases. This is due to increased compressor capacity causing the suction pressure to decrease, therebydeveloping a higher specific volume of the suction line gas. Assume, for example, that the ambient temperature drops to 60 F. At that temperature refrigerant R-22 has a gas-toliquid ratio of 52 to l, 1 cubic foot of liquid generating 52 cubic feet of refrigerant gas. Once again the system will be unbalanced but this time there is effectively more gas than the transformers ratio of 45 to 1 ordinarily allows. Hence, in order to pump the increased volume of gas through the secondary more energy will be needed in the secondary and must be delivered from the primary. The liquid flow in the primary must therefore increase so that the energy demand is satisfied, and

this gives rise to an increase in operating speed. With more refrigerant flowing to the evaporator, a portion thereof begins to flood out of the evaporator and through the secondary. The refrigerant in the secondary will now be a mixture of gas and liquid and its specific volume will decrease until the system rebalances under conditions of sufficient liquid flood through to satisfy the transformers requirement of 45 cubic feet of refrigerant flowing through the secondary for every l cubic foot of refrigerant in the primary. Thus, in response to a lower outside ambient the transformer settles down or stabilizes at an operating speed faster than that prevailing under the rated conditions. More pounds per minute of refrigerant will be fed to the evaporator and it will be filled at all times with boiling liquid.

it is significant that at outdoor temperatures lower than F the fluid transformer tends to overfeed the evaporator while all other conventional devices such as capillary tubes and expansion valves will underfeed. The liquid overfeed of the fluid transformer is eliminated by a simple control device to be discussed later.

The described mode of operation, in which the primary circuit drives the secondary circuit, occurs so long as the outside ambient remains sufficiently high to hold the condenser head pressure above a predetermined minimum level appropriate to retain a higher pressure at the primary inlet than at the primary outlet. For example, the refrigeration system of FIG. 1 could be designed so that the required pressure differential across the primary circuit would exist at all outdoor temperatures exceeding 25 F. At lower ambient temperatures the condensing pressure will be below the required minimum, in which case the pressure at the primary outlet will be greater than that at the primary inlet. The system must now function in accordance with an entirely different mode of operation as there is no longer any expanding liquid in the primary to impart energy to the secondary. Furthermore, it is necessary to force or pump the liquid from the lower condensing pressure to the higher pressure at the primary outlet. Switching from one mode to the other is automatically done by the fluid transformer itself anytime the ambient conditions so warrant.

To explain, inasmuch as the condensing pressure during low ambients sets the lowest pressure, the pressure at which evaporation occurs will constitute the highest pressure. The suction gas entering the secondary inlet is thus established at the highest pressure in the system and this permits the secondary circuit to take over as the prime mover in the transformer. Stated differently, the nominal primary and secondary circuits interchange roles with the secondary now driving the primary. The gaseous refrigerant expands in the secondary and energy is transfered from the secondary to the primary to achieve pumping of the liquid refrigerant from the primary inlet to the outlet.

The operation of the system during low ambient conditions is shown by the pressure-enthalpy diagram of FIG. 5. With the condenser pressure below the evaporator pressure, free cooling is obtained as the cycle will be self sustaining without requiring operation of the compressor. Refrigerant boiled in the evaporator flows through the secondary circuit, expanding to the condenser pressure. The pressure drop in the secondary is denoted AP, in H6. 5. After leaving the secondary the refrigerant flows through the non-running compressor and then through the condenser. The refrigerant is then pumped in the primary to the higher evaporator pressure to complete the cycle. The pressure differential labeled AP indicates the differential through which the liquid refrigerant must be pumped.

The free cooling provided during low ambient conditions is not possible in previous arrangements following the conventional technique of artificially holding the condensing pressure up so that there is always a relatively high liquid line pressure regardless of the outside temperature. The advantages of free cooling during low ambients, when compressor operation is not needed, and the improved coefficient of performance during middle ambients, when the compressor is running, have not been achieved heretofore. There will be no need to use outdoor air for cooling during low ambient conditions and this eliminates all of the associated mechanical apparatus required in prior systems such as dampers, damper motors, controls, outdoor air ductwork, weather proof grilles, power exhaust fans and motors, filters, weather sealing, etc. In addition, in the absence of outdoor air no high humidity, odors, dirt, etc. will be introduced.

Of course, regardless of which flow circuit powers the transformer refrigerant will be automatically metered to the evaporator, by the action of the transformer, to maintain it full of boiling liquid and to produce a suction line to liquid line volume ratio of 45 to 1. For example, if the outside temperature drops to F the gas-to-liquid ratio will normally be 60 to l causing the system to unbalance. At that outdoor temperature the system will operate in accordance with its low ambient mode and the gas in the secondary pumps the liquid through the primary. The over abundance of gas flow in the secondary causes the fluid transformer to increase its speed, as a result of which the liquid flow through the primary to the evaporator increases. Part of the liquid is flooded through into the secondary, the specific volume of the mixture in the secondary decreasing so the system is balanced once again.

The operation of the fluid transformer is also governed and effectively controlled by the amount of evaporator heat loading, the flow rate of the refrigerant in each of the flow circuits varying inversely with that loading in order to properly meter and regulate the refrigerant flow in the vapor compression cycle in response to different heat loads. As the evaporator load increases, more liquid is boiled into gas and additional superheat is added to the gas before it arrives at the evaporators outlet. The suction gas temperature, and consequently its pressure, increase and this results in a decrease of the specific volume of the gas flowing through the secondary circuit. The gas-to-liquid ratio now becomes less than the required 45 to l and the transformer's operating speed decreases in the manner previously described, thereby feeding less liquid to the evaporator to bring the evaporator pressure back to a value necessary to provide a gas specific volume of 45 times that of the liquid. Hence, the system does not allow higher than normal or desired suction pressure due to the evaporator heat loading in the same manner as the fluid transformer limits the evaporator pressure in the presence of high condenser head pressures.

On a decreasing evaporator heat load, the suction pressure tends to drop and this increases the specific volume of the suction gas. With the gas-to-liquid ratio now being higher than 45 to l, the fluid transformer increases its speed and overfeeds the evaporator. Sufficient overfeed will occur to return the volume ratio back to 45 to l in the same fashion that takes place under low head pressure conditions. It is thus seen that the transformers operating speed varies inversely with both head pressure and evaporator loading.

In response to compressor shut down, flow in the secondary ceases immediately and the transformer stops running. A blockage is therefore provided in the liquid line and this prevents overfeeding of the evaporator, leaving it comparatively dry. Conventional expansion devices such as capillary tubes continue to feed as long as there is a pressure difference. An expansion valve will overfeed because of its slow response time as a result of which the evaporator will be left relatively full of refrigerant to satisfy its superheat setting.

At start-up, the fluid transformer begins to feed the evaporator immediately. Conventional devices require a build-up of the head pressure to provide the pressure differential necessary to supply the evaporator. Consequently, the fluid transformer eliminates the evaporator pressure dip experienced with conventional devices and provides cooling capacity immediately.

As thus far described, under conditions of low head pressure or low evaporator pressure some liquid flood back occurs. To modulate liquid flow in the presence of these conditions, hot discharge gas from the compressor is injected at the primary inlet under the control of a reverse acting thermostatic expansion valve 18. This hot gas occupies some of the displacement in the primary so that while the total volume flow rates are still at the ratio of 45 gas side to 1 liquid side, a portion of the liquid side displacement is spoiled" with the hot discharge gas so that the true ratio of suction gas flow to actual liquid flow.can be modulated in an upward direction with increased hot gas. At the rated outside temperature of F no hot gas is required. Preferably, the system would be designed so that at 0 F outside air a 20 percent overfeed would occur with no control. This means that 25 percent of the desired liquid flow volume would have to be supplied by hot gas. Since under these conditions the volume ratio of hot gas to liquid is approximately 55 to 1, only 0.45 percent 1/55 times 25 percent) of the total flow rate is sacrificed to achieve control. This is insignificant in the overall operation and even though under normal conditions evaporator freeze-up will not take place until the outdoor ambient is below -20 F, the small amount of hot gas bypass does provide a slight edge in favor of preventing freeze-up.

Reverse acting expansion valve 18 functions in well known manner in response to evaporator superheat, having a valve which opens to an extent inversely proportional to the amount of superheat present in the suction gas. Specifically, there are two controls that combine to set the size of the valve opening. Thermal sensing bulb 19 responds to the suction line temperature and exerts, on one side of a diaphragm in device 18, a force representing that temperature. A pressure sensing tube 21 is in direct communication with the suction line and exerts on the other side of the diaphragm a force proportional to the suction line pressure and this in turn is proportional to the saturation temperature of the refrigerant. The difference between the two forces effectively corresponds to the superheat and provides a net force that determines the size of the valve opening, the greater the superheat the smaller the opening. As evaporator superheat rises, the quantity of hot discharge gas introduced into the liquid flow to the primary inlet decreases, and conversely, decreased superheat results in less gas being added to the liquid at the primary inlet. As a result, the quantity of liquid flow to the evaporator increases on a superheat increase and decreases on a superheat decrease, providing constant evaporator superheat (in the desired amount) under all operating conditions.

Expansion valve 18 is effectively connected across condenser 12, as a consequence of which the pressure drop across the valve equals the condenser pressure drop. The required valve capacity is minimum at high condenser pressures when the condenser pressure drop is also minimum, while the required valve capacity is maximum at low condenser pressures at which time the condenser pressure drop is maximum. This means that the valve has sufficient capacity under all operating conditions to maintain the system under positive control, the system stability thus being exceptional as compared to that obtained in a conventional system where a thermostatic expansion valve directly meters liquid refrigerant.

Consideration will now be given to the preferred construction of the fluid transformer shown in each of FIGS. 2 and 3. In general, the transformer comprises two similar fluid actuated piston systems operating at the same speed (the transformers operating speed) but one-fourth of a cycle out of phase, each piston system automatically providing the intake and exhaust porting for the other. Collectively FIGS. 2 and 3 depict the transformer at four different steps of its operating cycle, the steps being spaced apart by one-fourth of a cycle.

More particularly, the transformer has a pair of cylinder structures 22 and 24 shown for convenience of illustration as two separate structures. Preferably, in practice structures 22 and 24 would be integrally related to each other in one composite cylinder block. Each structure is hollowed out to form a single large cavity which is effectively divided into five smaller cylindrical chambers having parallel axes. To explain, the cavity in structure 22 includes the large centrally located cylindrical chamber 26, the two small cylindrical chambers 27 and 28 on the right, and the two small cylindrical chambers 31 and 32 on the left, all four of chambers 27, 28, 31 and 32 having the identical crosssectional area. In like fashion, cylinder structure 24 is also hollowed out to provide a single large cavity containing the five smaller cylindrical chambers 34, 35, 36, 37 and 38.

The primary inlet couples to both chambers 31 and 36 via ports 39 and 41 respectively. The primary outlet communicates with chambers 32 and 35 by way of ports 42 and 43 respectively. Chambers 31 and 32 are interconnected via port 44, while chambers 35 and 36 communicate through port 45. The secondary inlet connects to two different sections of chamber 38 by means of ports 46 and 47, and communicates through ports 48 and 49 to two different points along chamber 27. The secondary outlet is connected via ports 51 and 52 to two different sections of chamber 37 and through ports 53 and 54 to two different parts of chamber 28. Chambers 27 and 28 are coupled together through port 55, while chambers 37 and 38 are interconnected via port 56. Chambers 31 and 38 are cross coupled through ports 57 and 58 and the line or conduit connected therebetween. Similarly, ports 58 and 59 and the intermediate fluid line permits cross coupling of chambers 28 and 35.

In addition, chamber 26 is connected to chambers 37 and 38 and chamber 34 connects to chambers 27 and 28. These interconnections are made by means of a series of ports whose axes are perpendicular to the crosssectional view shown in each of FIGS. 2 and 3; consequently those ports are illustrated in circular configuration. Each cylinder structure 22 and 24 has eight such ports designated by the reference numbers 61-68. They have been given the same reference numerals in both cylinder structures since each pair of correspondingly numbered ports actually constitutes only a single port. As mentioned, structures 22 and 24 preferably would be constructed as a single composite cylinder block. By forming the cylinder block so that structures 22 and 24 are effectively back-to-back, port 61 of structure 22 may be aligned and made coincident with port 61 of structures 24, while at the same time each of the other seven pairs of correspondingly numbered ports may likewise be aligned and made coincident. Hence, in reality there need be only a total of eight ports in the composite cylinder block.

Reciprocally movable through the large cavity of each structure 22, 24, and through a two-stroke cycle, is an associated free piston mechanism which comprises five separate cylindrical pistons all mechanically fixed to each other, each movable within a respective one of the five chambers of the cavity. Specifically, free piston mechanism has five different portions 71-75 forming separate pistons sized for movement within respective ones of chambers 26, 31, 32, 27 and 28. Piston 72 is further divided into a full diameter portion 72a to provide a land for blocking port 57, a reduced diameter portion 72c for intercoupling ports 57 and 44, and a full diameter portion 72b. Piston 74 is broken up into three full diameter portions 74a, 74b and 740, with a pair of intermediate reduced diameter portions 74d and 74e, portions 74b providing lands for closing ports 48, 49, 65 and 67, portion 74a effecting intercoupling of ports 48 and 65, and portion 74e facilitating intercoupling of ports 49 and 67. Piston 75 has a pair of full diameter portions 75 a and 75b between which is a reduced diameter portion 75c, portion 75a causing blocking of ports 53 and 66, portion 75c effecting intercoupling of ports 53 and 66 and also of ports 54 and 68.

In like manner free piston mechanism 80, whose shape is similar to that of piston mechanism 70, combines five different pistons 81-85 each of which is reciprocally movable within respective ones of chambers 34, 35, 36, 37 and 38. Piston 82 has a pair of full diameter portions 820 and 82b separated by a reduced diameter portion 820. Piston 84 is divided into three full diameter portions 84a, 84b and 840 and a pair of reduced diameter portions 84d and 84e. Piston 85 has full diameter portions 85a and 85b and an intermediate reduced diameter portion 850. The various portions of pistons 82, 84 and 85 effect opening and closing of ports in structure 24 in the same manner as described above in connection with piston mechanism 70.

Of course, in a practical construction of the fluid transformer several piston rings would be employed. They are especially desirable since both the primary and secondary fluid flow through different parts of each of the two cavities and it is important to isolate the primary and secondary flow circuits so that there is no leakage where they interface. The piston rings have not been illustrated in order to avoid encumbering the drawings.

It will be recalled that the fluid transformer has two different operating modes; in one (the high ambient mode) the nominal primary drives the nominal secondary and in the other mode (the low ambient mode) the secondary powers the primary. In the high ambient mode the primary effectively functions as a fluid actuated expansion engine and the secondary acts as a fluid pump, the liquid serving as the driving fluid for powering each of free piston mechanisms 70, 80, through its two-stroke cycle and the gas serving as the driven or pumped fluid. In the low ambient mode of operation the functions are reversed.

The operation of the transformer will initially be described under high ambient conditions at which time the primary circuit serves as the prime mover and the primary inlet and outlet are established at the high side and low side pressures respectively. For convenience, the pressure of the liquid refrigerant at the primary inlet may be referred to as the drive pressure, while the lower pressure at which the refrigerant is discharged at the primary outlet may be called the exhaust pressure. In FIG. 2 piston mechanism 70 is shown in full line construction halfway through one of its strokes and moving to the right as indicated by arrow 91. Piston mechanism 80 lags mechanism 70 by one-quarter of a cycle so it is shown in full line construction in FIG. 2 in its extreme position on the left and starting to move to the right as indicated by arrow 92. For convenience, mechanism 70 may be referred to as the leading piston mechanism and mechanism 80 as the lagging piston mechanism. Each of the piston mechanisms is also shown in FIG. 2 in dashed construction one-fourth of a cycle later, mechanism 70 therefore being at its rightmost position while mechanism 80 is midway through its travel or stroke and heading toward the right. One-quarter of a cycle still later mechanism 70 is halfway through its other stroke and moving to the left as shown in full line construction in FIG. 3 and as is indicated by arrow 93. At that same time mechanism 80, lagging by one-fourth of a cycle, will be established at its rightmost position, as shown in full line construction in FIG. 3, and will be starting to move to the left as indicated by arrow 94. Finally, one-fourth of a cycle later mechanism 70 will be at its extreme left position while mechanism 80 will be centered and travelling to the left, as illustrated by the dashed forms of the piston mechanisms in FIG. 3. Piston mechanisms 70 and 80, one-quarter of a cycle subsequently, will return to their positions shown in full line construction in FIG. 2.

The manner in which the piston mechanisms are cyclically actuated through the illustrated positions in response to liquid refrigerant flowing in the primary circuit will now be described. Movement of each of the piston mechanisms to the right is effected by establishing the fluid at the left end or side of the mechanism at the higher drive pressure, while at the same time the fluid at the right end is established at the much lower exhaust pressure. Conversely, movement of each piston mechanism to the left is caused by establishing the fluid on the right side at the drive pressure and the fluid on the left side at the exhaust pressure. The two ends of each piston mechanism are alternately established at drive and exhaust pressure by valving controlled by the other piston mechanism. As will be made apparent the primary flow circuit of the fluid transformer may be thought of as comprising means, which includes a portion of piston mechanism 80, for providing intake and exhaust valving of the large cavity in cylinder structure 22 to effect, in response to the driving fluid, double acting reciprocating movement of piston mechanism through its two-stroke cycle. In addition, the primary circuit may be considered as also comprising means, which includes a portion of piston mechanism 70, for providing intake and exhaust valving of the large cavity in cylinder structure 24 to effect, in response to the liquid refrigerant, double acting reciprocating movement of piston mechanism through its two-stroke cycle. To explain, consider initially the action of the fluid transformer as piston mechanisms 70 and 80 travel between their positions shown in FIG. 2, namely as mechanism 70 moves from its center to its rightmost position while mechanism 80 shifts from its leftmost to its center position. During that quarter of a cycle, the driving fluid flows from the primary inlet through port 39 and into chamber 31 and thence through port 57 (the land provided by the full diameter portion 72a will be moving to the right) to chambers 37 and 38, those chambers being intercoupled via port 56 throughout the entire cycle. The left ends of pistons 72, 84 and will therefore be established at the drive pressure. Meanwhile, chambers 27 and 28 (intercoupled via port 55 at all times) are coupled through ports 58 and 59, chamber 35 and port 43 to the primary outlet. Hence, the right ends of pistons 74, 75 and 82 will be at the lower exhaust pressure. It is also to be noted that chamber 36 communicates through port 41 to the primary inlet while cavity 32 couples through port 42 to the primary outlet. The net pressure at the left end of each piston mechanism, however, will be the drive pressure while the net pressure on their right ends will be the exhaust pressure.

For a complete understanding of the operation, it is necessary to consider the volume of liquid refrigerant flowing at any given instant into the primary inlet and also out of the primary outlet. As the piston mechanisms are traveling to the right and through their quarter cycles shown in FIG. 2, the volume of liquid flowing into chamber 31 per unit of time may be designated AV. Since all eight of the small pistons have the identical cross sectional area, each of chambers 37 and 38 effectively displaces liquid in the amount of AV from the primary inlet and through ports 39, 57 and 58. Hence, a total of three AV is being drawn from the primary inlet by chambers 31, 37 and 38. As piston 73 moves to the right liquid in the amount AV is being sucked or robbed from the discharge or primary outlet. As a result, a net amount of three AV is being drawn from the primary inlet at the left side of the fluid transformer and one AV is being robbed from the discharge. On the right side of the transformer, each of pistons 74, 75 and 82 is pumping or pushing liquid in the amount AV toward the primary outlet. Since piston 83 is moving to the right, and thus working against the primary inlet, the net effect on the right side of the transformer is to deliver the amount three AV to the outlet and one AV to the inlet. As a consequence, by combining the left and right ends at any given instant the same amount of fluid (two AV) is being sucked in from the primary inlet as is being discharged out at the primary outlet.

When leading mechanism 70 reaches its extreme right position its motion will be reversed by valving controlled by mechanism 80. Since lagging mechanism 80 will now be centered but still traveling to the right, the land provided by full diameter portion 82a momentarily blocks port 59 (as shown in dashed construction in FIG. 2) to disconnect chambers 27 and 28 from the discharge outlet. As mechanism 80 continues its rightward movement chambers 27 and 28 will communicate with the primary inlet via ports 58, 59, 45 and 41, chamber 36 and the annular portion of chamber 35 surrounding reduced diameter section 820. The right ends of both pistons 74 and 75 will thus be established at the drive pressure to propel mechanism 70 to the left. Driving fluid will still be supplied to chamber 31. However, since both chambers 27 and 28 receive driving fluid the net effect is to maintain the right end of mechanism 70 at the drive pressure and the left side at the exhaust pressure. In other words, the force introduced by the driving fluid in chamber 31 will be offset or canceled by the opposing force presented by the driving fluid in one of chambers 27, 28.

During the quarter cycle in which leading mechanism 70 moves to the left from its far right position to its center position shown in full line construction in FIG. 3, during which time mechanism 80 continues its movement to the right to its rightmost position shown in full line construction in FIG. 3, it can be shown that at any given instant liquid in the amount two AV will be drawn from the inlet and the same amount will be exhausted to the outlet. More particularly, on the left side of the transformer chambers 37 and 38 together displace in from the primary inlet the amount two AV, as those chambers communicate with the inlet through ports 58, 57 and 39 and chamber 31. However, since piston 72 is moving to the left and against the source of driving fluid, chamber 31 in effect discharges driving fluid in the amount AV. Actually, what this means is that one AV of the fluid supplied to chambers 37 and 38 will be pushed into those chambers by piston 72. The net amount therefore drawn from the inlet at the left side of the transformer will be 2AV-AV or one AV. At the same time, piston 73 will be displacing liquid in the amount AV out of chamber 32 and into the discharge outlet. On the right side of the transformer, chambers 27 and 28 will be displacing the amount two AV from the primary inlet and around the reduced diameter portion 820. Piston 83, however, moving to the right will effectively subtract one AV from the liquid displaced in from the primary inlet so that the net amount drawn will be one AV. Meanwhile, piston 82 will be pumping an amount one AV out of chamber 35 into the primary outlet. Thus, by adding the amounts displaced from the inlet by both the left and right sides and by adding the amounts delivered to the outlet by both sides, it is found that at any given instant the amount two AV of liquid refrigerant is displaced in from the primary inlet and the same amount is exhausted at the primary outlet.

When leading mechanism 70 reaches its center position while simultaneously lagging mechanism 80 reaches its extreme right position, the intake and exhaust porting for mechanism 80 will be reversed by mechanism 70, as a consequence of which mechanism 80 reverses its motion and begins to move to the left. Specifically, at the instant mechanism 80 reaches its far right position (shown in full line construction in FIG. 3) full diameter portion 72a blocks port 57 and cuts off the supply of driving fluid to chambers 37 and 38. The liquid flowing into chamber 36 from the primary inlet will now be effective to propel mechanism 80 to the left. As that is taking place, mechanism continues its leftward movement toward its extreme left position (shown in dashed construction in FIG. 3) and this permits chambers 37 and 38 to communicate with the primary outlet through a path which includes ports 58, 57, 44 and 42, chamber 32 and the annular section of chamber 31 surrounding reduced diameter portion 72c. An analysis of the liquid flow into and out of both sides of the transformer, for the one quarter cycle during which mechanism 70 shifts from its center to far left position and lagging mechanism travels from its far right to its center position, will reveal that piston 72 forces an amount AV towards the primary inlet while each of pistons 73, 84 and 85 exhausts an amount AV toward the primary outlet. On the left side there is therefore one AV displaced toward the inlet and three AV displaced toward the outlet. On the right side, an amount AV flows into each of chambers 27, 28 and 36 from the primary inlet while leftward moving piston 82 effectively robs an amount AV from the primary outlet. On the right, therefore, a total of three AV is displaced in from the inlet while one AV is taken from the outlet. Combining the flows on both sides, however, it is seen that at any given instant an amount two AV is drawn from the primary inlet and an amount two'AV is exhausted at the primary outlet.

To complete the analysis of the primary flow circuit, when leading mechanism 70 reaches its far left position and mechanism 80 reaches its center position, mechanism 80 once again reverses the intake and exhaust porting for piston mechanism 70 in order to effectively establish the left and right ends of the cavity of cylinder structure 22 at drive and exhaust pressures, respectively, thereby to reverse the movement of mechanism 70. Specifically, at the instant mechanism 70 reaches its far left position the land provided by full diameter portion 82a blocks port 59 to cutoff the supply of driving fluid to chambers 27 and 28, as a result of which the fluid in chamber 31 (which is at drive pressure) forces mechanism 70 to the right. While this is occurring mechanism 80 is moving leftward from its center position to unblock port 59 so that chambers 27 and 28 will now communicate with the primary outlet via ports 58, 59 and 43. A consideration of the intake and exhaust flow during the one fourth cycle in which mechanism 70 moves from its far left to its center position and mechanism 80 shifts from its center to its far left position, reveals that each of chambers 31 and 36 draws an amount AV from the primary inlet, each of chambers 37, 38, 27 and 28 delivers an amount AV toward the primary outlet, and each of chambers 32 and 35 robs an amount A V from the primary outlet. Adding of the amounts indicates that at any given instant liquid refrigerant in the amount two AV is drawn from the primary inlet and the same amount is exhausted at the primary outlet.

The driven fluid, namely the refrigerant gas, is pumped through the secondary flow circuit by the reciprocating or oscillating action of piston mechaism 70 and 80. As large pistons 71 and 81 travel within their respective chambers 26 and 34, gas is drawn in at the secondary inlet and discharged at the secondary outlet. The intake and exhaust valving of chamber 26 is controlled by pistons 84 and 85 of piston mechanism 80, while the intake and exhaust porting of chamber 34 is controlled by pistons 74 and 75 of piston mechanism 70. The two cylinder structures are cross valved in the secondary circuit in much the same manner as the cross valving in the primary circuit that causes the reciprocating movement of the eight small pistons. During each stroke of each of pistons 71 and 81, one end of the associated chamber 26, 34 communicates with the secondary outlet while the other end couples to the secondary inlet.

More particularly, during the quarter cycle in which leading mechanism 70 is traveling from its center to its far right position and lagging mechanism 80 is shifting from its far left to its center position (see FIG. 2), gas is fed from the secondary inlet through port 46 and around reduced diameter portion 85c to supply port 61 and thence to the left end of chamber 26, exhaust port 62 being blocked at that time by full diameter portion 8412. Meanwhile, supply port 63 is closed by full diameter portion 85a but exhaust port 64 will be opened to communicate the right end of that chamber to the secondary outlet over a path comprising port 52 and the annular section of chamber 37 encompassing reduced diameter portion 84d. As a consequence, as piston 71 moves toward its far right position refrigerant gas from the secondary inlet is drawn into chamber 26 on the left side of piston 71 and is pumped out from the right side of the piston to the secondary outlet.

Concurrently, pistons 74 and 75 valve chamber 34 in like fashion so that the left end of the chamber is coupled to the secondary inlet and its right side to the secondary outlet. Specifically, supply port 65 is opened by reduced diameter portion 74d and communicates with the secondary inlet via port 48, exhaust port 66 being closed at that time by full diameter portion 75a. Meanwhile, discharge port 68 is opened by reduced diameter portion 750 so that the right end of chamber 34 communicates through port 54 to the secondary outlet. Supply port 67 is blocked at that time by full diameter portion 74b.

When mechanism 70 reaches its far right position and begins to return toward its center position mechanism 80 will still be traveling rightward but will have crossed its center position. During that quarter cycle, valving controlled by portions of pistons 84 and 85 reverse the intake and exhaust for chamber 26 in order that the leftward moving piston 71 now pumps gas from the left end of chamber 26 into the secondary outlet and draws gas into that chamber at the right end. Particularly, port 62 is opened to the secondary outlet by means of reduced diameter portion 84c and port 51. Port 61 is closed by full diameter portion 85b, port 63 is opened to the secondary inlet via port 47 and the section of chamber 38 surrounding reduced diameter 85c, and port 64 is closed by full diameter portion 84b.

In like manner, when mechanism 80 reaches the end of its rightward moving stroke pistons 74 and 75 reverse the intake and exhaust valving for chamber 34 so that refrigerant gas will now be discharged to the secondary outlet from the left side of chamber 34 and drawn in from the secondary inlet at the right end as piston mechanism travels through its leftward moving stroke.

Under low ambient conditions when the condensing pressure at the primary inlet represents the lowest pressure of the system, the fluid in the primary flow circuit will be unable to provide the necessary power to effect reciprocating movement of each piston mechanism. However, since the intake and exhaust valving for the secondary circuit functions in the same manner as that for the primary circuit, the relatively high pressure gas at the secondary inlet serves as the driving or powering fluid to effect reciprocating movement of pistons 71 and 81 and consequently of mechanisms 70 and 80. Hence, during the low ambient operating mode refrigerant will be pumped from the primary inlet to the primary outlet in the same way as gas is pumped through the secondary during the high ambient mode of operation.

It will be observed that exhaust ports 62, 64, 66 and 68 are offset from their associated intake ports. By such positioning, as each of pistons 71 and 81 is powered by the refrigerant to the end of a stroke the piston itself covers the previously open exhaust port thereby providing a bounce chamber for reversing the motion of the associated piston mechanism.

The volumetric flow ratio of the fluids in the primary and secondary circuits is determined by the total volume displaced by the pistons in each of those circuits during the same interval, such as during one complete stroke of either one of the piston mechanisms. Assuming that the system has stabilized in that there is no change in evaporator loading or outside ambient, the primary volume displacement would therefore be equal to the total volume of fluid emerging at the primary outlet during the time interval required for mechanism 70, for example, to travel through one complete stroke. Similarly, the secondaryvolume displacement will equal the total volume of refrigerant gas discharged at the secondary outlet during one complete stroke of mechanism 70. Of course, in accordance with the description of the illustrated refrigeration system of FIG. 1 the secondary volume displacement should be approximately 45 times the primary volume displacement when refrigerant R22 is used. The primary and secondary circuits have not been so proportioned in the drawings of FIGS. 2 and 3 in order that the primary circuit may be shown in greater detail.

Actually, the same results as achieved in the illustrated refrigeration system may also be realized with a fluid transformer having a substantially smaller ratio of its secondary volume displacement versus its primary volume displacement. For example, a transformer with a ratio of IS to I could be used. All of the liquid refrigerant from the condenser outlet would flow through the primary circuit but only one-third of the refrigerant gas from the evaporator would be channeled through the secondary. This may easily be done by providing the evaporator with three different branches or coils and by connecting the primary outlet to a distributor having three outlets respectively connected to the inlets of the three branches. The outlets of two of the branches, representing two-thirds of the evaporator, are connected directly to the suction inlet of the compressor. The outlet of the other branch coil, constituting the remaining one-third of the evaporator, connects to the secondary inlet of the transformer. The secondary outlet connects to the suction inlet as in FIG. 1. With such an arrangement, liquid will be metered through the distributor to the evaporator on the basis of one forty-fifth times the volume flow of refrigerant gas received at the suction inlet of the compressor. There is a substantial pressure drop in distributor so it would serve as part of the expansion device for the compression cycle. The modified system could be constructed so that under high ambient conditions the pressure of the refrigerant would drop from, for example, around 300 psig at the primary inlet to about 120 psig at the primary outlet, a further reduction to around 70 psig taking place in the distributor. The pressure-enthalpy diagrams of FIGS. 4 and 5 would not, of course, apply to the modified refrigeration system. A major change in the low ambient diagram is necessary since the liquid would have to be pumped through the primary circuit from the relatively low condenser pressure (for example, around 60 psig) to the pressure (120 psig) required at the inlet of the distributor. In the FIG. 1 system the liquid is pumped in the primary, during low ambients, over a much smaller pressure differential.

As mentioned, in accordance with the invention there are many different ways in which a fluid transformer may be incorporated in a refrigeration system to achieve a variety of advantageous results. As another example, such a transformer can be used in the refrigeration system of the well known type wherein the evaporator is overfed into an accumulator and the overfeed is recirculated through the evaporator by means of an electrically powered pump. This improves the heat transfer property of the evaporator to maximize the heat absorbed from the load. The pump could be replaced by a fluid transformer, the driving power being supplied by the refrigerant gas.

Applicant has therefore provided a unique flow control arrangement that is very versatile and may be employed in a refrigeration system in a variety of different configurations. It features a fluid transformer which uses refrigerant flowing in one phase of the compression cycle to drive refrigerant through another phase of the cycle or through another portion of the same phase.

Certain features described in the present application are disclosed and claimed in copending application Ser. No. 1 15,507, filed concurrently herewith in the name of the present applicant, and assigned to the present assignee.

While a particular embodiment of the invention has been shown and described, modifications may be made, and it is intended in the appended claims to cover all such modifications as may fall within the true spirit and scope of the invention.

I claim:

1. A refrigeration system in which refrigerant flows through a vapor compression refrigeration cycle, comprising:

fluid transformer means for utilizing the refrigerant flowing in one portion of the cycle to control the refrigerant in another portion of the cycle and having a primary flow circuit, with an inlet and an outlet, connected in series with said one portion of the cycle and a secondary flow circuit, with an inlet and an outlet, connected in series with said other portion of the cycle,

said fluid transformer functioning, when the pressure at the primary inlet is higher than that at the primary outlet, in one operating mode wherein the refrigerant flowing through said primary circuit effects pumping of the refrigerant through said secondary circuit and functioning, when the primary inlet pressure is less than the primary outlet pressure, in another operating mode wherein the refrigerant flow in the secondary circuit causes pumping of the refrigerant through said primary circuit,

and in which said fluid transformer means maintains a constant ratio of the volume of refrigerant flowing through said secondary circuit per unit of time relative to the volume of refrigerant passing through said primary circuit per same unit of time, a change in flow rate of the refrigerant in either one of said circuits resulting in a corresponding flow rate change in the other circuit in accordance with that ratio.

2. A refrigeration system according to claim 1 in which each of said flow circuits has at least one piston member reciprocally movable at a speed varying directly with the volume of refrigerant flowing through the circuit per unit of time, and in which each piston member of said first circuit is mechanically fixed to a piston member of said second circuit to provide, for said flow circuits, identical speeds and equal to the transformers operating speed.

3. A refrigeration system according to claim 1 in which the refrigerant is in its liquid state in said one portion of said cycle and in its gaseous state in said other portion, during said one operating mode the liquid refrigerant expanding in said primary circuit and the gaseous refrigerant compressing in said secondary circuit, said primary and secondary circuits interchanging roles during said other operating mode such that the gaseous refrigerant in said secondary circuit may be utilized to effect pumping, and to increase the pressure, of the liquid refrigerant in said primary circuit.

4. A refrigeration system according to claim 1 in which the vapor compression refrigeration cycle includes, in series flow relationship, an evaporator, a compressor, a condensor and an expansion device, the refrigerant flowing in its gaseous state from said evaporator to said condensor and in its liquid state from said condensor to said expansion device, wherein at least part of the gaseous refrigerant leaving said evaporator flows through said secondary circuit before it reaches the inlet of said compressor, in which the liquid refrigerant leaving said condensor flows through said primary circuit before it reaches the inlet of said evaporator, said primary flow circuit also serving as at least part of said expansion device, in which the operation of said fluid transformer means is governed and effectively controlled by the condenser head pressure, the liquid refrigerant expanding in said primary circuit and imparting energy to said secondary circuit only when the condenser head pressure exceeds a predetermined minimum level sufficient to establish the primary inlet at a higher pressure than that at the primary outlet, condenser pressure below that minimum level causing said secondary circuit to take over as the prime mover of said fluid transformer with the result that the gaseous refrigerant expands in said secondary circuit and energy is transferred from said secondary to said primary circuit to achieve pumping of the liquid refrigerant and an increase in pressure from the primary inlet to the primary outlet.

5. A refrigeration system according to claim 4 in which the operation of said fluid transformer means is governed and effectively controlled by the amount of evaporator heat loading, the flow rate of the refrigerant in each of said flow circuits varying inversely with evaporation loading in order to properly meter and regulate the refrigerant flow in the vapor compression cycle in response to different evaporator heat loads.

6. A refrigeration system according to claim 4 in which the pumping capacity of said compressor, in terms of volume per unit of time, varies inversely with the condenser head pressure, said fluid transformer means functioning as a metering device with the quantity of gaseous refrigerant, in terms of volume per same unit of time, flowing to said compressor by the effect of said secondary circuit being substantially equal to the compressor pumping capacity established by the condenser head pressure prevailing at the time, and with the quantity of liquid refrigerant flowing through said primary circuit having a fixed relationship to the quantity of refrigerant gas pumped by said compressor.

7. A refrigeration system according to claim 1 in which the volume of refrigerant flowing through said primary circuit per unit of time is directly proportional to the volume of refrigerant flowing through said secondary circuit per same unit of time and vice versa, any variation in the flow rate in either one of said flow circuits resulting in a corresponding flow rate change in the other flow circuit.

8. A refrigeration system according to claim 1 in which said fluid transformer means effects a constant volumetric flow ratio between the refrigerant passing through said two flow circuits under a wide range of condenser pressures and evaporator loading, said primary circuit including at least one chamber to provide a primary volume through which the refrigerant is displaced and said secondary circuit including at least one chamber to provide a secondary volume through which the refrigerant is displayed, the ratio of the primary and secondary volumes being equal to the flow ratio of the refrigerant in said two flow circuits.

9. A refrigeration system in which refrigerant flows through a vapor compression refrigeration cycle from an evaporator to a condenser and then back to the evaporator, the refrigerant being delivered in its gaseous state to the condenser inlet and leaving the condenser outlet in its liquid state, a metering arrangement for controlling the quantity of refrigerant flow from said evaporator to said condenser and from said condenser to said evaporator under a wide range of condenser head ressures, comprising:

fluid trans ormer means having a first flow circuit,

connected between the outlet of said condenser and said evaporator and receiving liquid refrigerant, and a second flow circuit connected between at least a portion of said evaporator and the inlet of said condenser and receiving gaseous refrigerant, condenser head pressures above a predetermined minimum level effecting expansion of the liquid refrigerant in said first flow circuit with a resulting transfer of energy to said second flow circuit to effect compression of the gaseous refrigerant in said second flow circuit, while condenser pressures below that predetermined minimum level cause expansion of the gaseous refrigerant in said second flow circuit with a consequent transfer of energy to said first circuit to achieve pumping of the liquid refrigerant to a higher pressure and in the direction of said evaporator. 

1. A refrigeration system in which refrigerant flows through a vapor compression refrigeration cycle, comprising: fluid transformer means for utilizing the refrigerant flowing in one portion of the cycle to control the refrigerant in another portion of the cycle and having a primary flow circuit, with an inlet and an outlet, connected in series with said one portion of the cycle and a secondary flow circuit, with an inlet and an outlet, connected in series with said other portion of the cycle, said fluid transformer functioning, when the pressure at the primary inlet is higher than that At the primary outlet, in one operating mode wherein the refrigerant flowing through said primary circuit effects pumping of the refrigerant through said secondary circuit and functioning, when the primary inlet pressure is less than the primary outlet pressure, in another operating mode wherein the refrigerant flow in the secondary circuit causes pumping of the refrigerant through said primary circuit, and in which said fluid transformer means maintains a constant ratio of the volume of refrigerant flowing through said secondary circuit per unit of time relative to the volume of refrigerant passing through said primary circuit per same unit of time, a change in flow rate of the refrigerant in either one of said circuits resulting in a corresponding flow rate change in the other circuit in accordance with that ratio.
 2. A refrigeration system according to claim 1 in which each of said flow circuits has at least one piston member reciprocally movable at a speed varying directly with the volume of refrigerant flowing through the circuit per unit of time, and in which each piston member of said first circuit is mechanically fixed to a piston member of said second circuit to provide, for said flow circuits, identical speeds and equal to the transformer''s operating speed.
 3. A refrigeration system according to claim 1 in which the refrigerant is in its liquid state in said one portion of said cycle and in its gaseous state in said other portion, during said one operating mode the liquid refrigerant expanding in said primary circuit and the gaseous refrigerant compressing in said secondary circuit, said primary and secondary circuits interchanging roles during said other operating mode such that the gaseous refrigerant in said secondary circuit may be utilized to effect pumping, and to increase the pressure, of the liquid refrigerant in said primary circuit.
 4. A refrigeration system according to claim 1 in which the vapor compression refrigeration cycle includes, in series flow relationship, an evaporator, a compressor, a condensor and an expansion device, the refrigerant flowing in its gaseous state from said evaporator to said condensor and in its liquid state from said condensor to said expansion device, wherein at least part of the gaseous refrigerant leaving said evaporator flows through said secondary circuit before it reaches the inlet of said compressor, in which the liquid refrigerant leaving said condensor flows through said primary circuit before it reaches the inlet of said evaporator, said primary flow circuit also serving as at least part of said expansion device, in which the operation of said fluid transformer means is governed and effectively controlled by the condenser head pressure, the liquid refrigerant expanding in said primary circuit and imparting energy to said secondary circuit only when the condenser head pressure exceeds a predetermined minimum level sufficient to establish the primary inlet at a higher pressure than that at the primary outlet, condenser pressure below that minimum level causing said secondary circuit to take over as the prime mover of said fluid transformer with the result that the gaseous refrigerant expands in said secondary circuit and energy is transferred from said secondary to said primary circuit to achieve pumping of the liquid refrigerant and an increase in pressure from the primary inlet to the primary outlet.
 5. A refrigeration system according to claim 4 in which the operation of said fluid transformer means is governed and effectively controlled by the amount of evaporator heat loading, the flow rate of the refrigerant in each of said flow circuits varying inversely with evaporation loading in order to properly meter and regulate the refrigerant flow in the vapor compression cycle in response to different evaporator heat loads.
 6. A refrigeration system according to claim 4 in which the pumping capacity of said compressor, in terms of volume per unit of time, varies inversely with the condenser head pressurE, said fluid transformer means functioning as a metering device with the quantity of gaseous refrigerant, in terms of volume per same unit of time, flowing to said compressor by the effect of said secondary circuit being substantially equal to the compressor pumping capacity established by the condenser head pressure prevailing at the time, and with the quantity of liquid refrigerant flowing through said primary circuit having a fixed relationship to the quantity of refrigerant gas pumped by said compressor.
 7. A refrigeration system according to claim 1 in which the volume of refrigerant flowing through said primary circuit per unit of time is directly proportional to the volume of refrigerant flowing through said secondary circuit per same unit of time and vice versa, any variation in the flow rate in either one of said flow circuits resulting in a corresponding flow rate change in the other flow circuit.
 8. A refrigeration system according to claim 1 in which said fluid transformer means effects a constant volumetric flow ratio between the refrigerant passing through said two flow circuits under a wide range of condenser pressures and evaporator loading, said primary circuit including at least one chamber to provide a primary volume through which the refrigerant is displaced and said secondary circuit including at least one chamber to provide a secondary volume through which the refrigerant is displayed, the ratio of the primary and secondary volumes being equal to the flow ratio of the refrigerant in said two flow circuits.
 9. A refrigeration system in which refrigerant flows through a vapor compression refrigeration cycle from an evaporator to a condenser and then back to the evaporator, the refrigerant being delivered in its gaseous state to the condenser inlet and leaving the condenser outlet in its liquid state, a metering arrangement for controlling the quantity of refrigerant flow from said evaporator to said condenser and from said condenser to said evaporator under a wide range of condenser head pressures, comprising: fluid transformer means having a first flow circuit, connected between the outlet of said condenser and said evaporator and receiving liquid refrigerant, and a second flow circuit connected between at least a portion of said evaporator and the inlet of said condenser and receiving gaseous refrigerant, condenser head pressures above a predetermined minimum level effecting expansion of the liquid refrigerant in said first flow circuit with a resulting transfer of energy to said second flow circuit to effect compression of the gaseous refrigerant in said second flow circuit, while condenser pressures below that predetermined minimum level cause expansion of the gaseous refrigerant in said second flow circuit with a consequent transfer of energy to said first circuit to achieve pumping of the liquid refrigerant to a higher pressure and in the direction of said evaporator. 